The present invention relates to vehicle transmissions, especially for heavy, on-as well as off-road vehicles, and more particularly to dual clutch transmissions with a range section that can be by-passed to allow it to be shifted without interrupting the power transfer.
The conventional stepped transmission for rear-wheel drive vehicles can be regarded as a robust, compact and cost-effective product. A typical example of such a lay-out is shown in FIG. 2 in DE10242823 A1. An input shaft is coaxial with a main (output) shaft and parallel with a countershaft. A gearwheel that is rotationally fixed on the input shaft meshes with a gearwheel that is rotationally fixed on the countershaft. Several pairs of meshing gearwheels are then located side by side. In each of these pairs, one gearwheel is coaxial with the countershaft; the other is coaxial with the main shaft. One of these gearwheels is rotationally fixed on its coaxial shaft. The other gearwheel, the loose gearwheel, is free to rotate relative to its coaxial shaft, but can be rotationally locked to the shaft by a mechanical tooth clutch. This construction is used both for manually shifted transmissions as well as for automated mechanically engaged transmissions, AMTs. Because of the mechanical tooth clutches, there will during the shift be an interruption of the power transfer between the engine and the driven wheels of the vehicle. Thus, this type of transmission is not powershifting.
For heavy road vehicles, e.g., heavy trucks and buses, a large number of gears is required. Often, the conventional stepped transmission construction is extended with a splitter and a range section. The basic principle of this is shown in FIG. 1 in U.S. Pat. No. 5,351,569. With the splitter, there are two alternate ways of transferring the power from the input shaft to the countershaft. This will double the number of gears. The range section can functionally be regarded as an additional gearbox that is connected in series with the ordinary, main transmission. There are two gears therein; a low range gear with a reduction ratio in the same order of magnitude as the largest reduction ratio of the main transmission, and a high range gear where there is no change in speed from the main transmission to the output shaft of the range section. Thus, a range section also has the ability to double the number of gears of the transmission.
Hence, with a splitter and range construction of a transmission, a large number of gears, e.g., 12 or 16, can be embodied using a limited number of gearwheels. Furthermore, with a range section, large reduction ratios can easily be embodied. Such ratios are required for start-off from rest and low-speed manoeuvres of a heavy vehicle. Still, there will be power transfer interruptions at gear shifts. Most heavy vehicles have turbo-charged diesel engines that have slow dynamic response. Then, the power interruptions at gear shifts will have an unfavourable impact on the performance of the vehicle. This is especially the case at off-road driving.
Powershifting transmissions are available for heavy road vehicles. Such transmissions are, in general, based on planetary gear trains and wet multi-plate clutches, e.g., as in EP0073102A2. That makes them expensive, and they have considerable power losses. Thereby, they are not suited for vehicles of the future, where fuel consumption will be more important than ever.
Hence, a cost-efficient transmission with low power losses and no power interruptions at gear shifts would be desirable for heavy on- and off-road vehicles.
Moreover, it would be of further advantage if it could be similar in size to conventional transmissions for facilitating the installation in the vehicle chassis.
Dual clutch transmissions are an interesting cross-breed between powershifting planetary transmissions and conventional stepped transmissions with power interruption at gear shifts. In principle, a dual clutch transmission has two input shafts, each connectable with a friction clutch to the output of the engine. Functionally, this is equivalent to having two conventional transmissions in parallel, i.e., two parallel sub-transmissions, and using one at a time for power transfer. The sub-transmission that is not used, idling, for the time being, can have a gear engaged and prepared for a subsequent shift. This shift is carried out by simultaneously disengaging the friction clutch of the previously used sub-transmission and engaging the friction clutch of the previously idling sub-transmission.
When properly designed, dual clutch transmissions have a potential of providing powershifts at a reasonable production cost and low power losses. This is due to the fact that the rotating parts, i.e., gearwheels, shafts and tooth clutches, are similar to those in conventional stepped transmissions. This, furthermore, enables the use of the same production equipment. So, it makes sense to produce dual clutch transmissions in the same facilities as used for conventional stepped transmissions.
Dual clutch transmissions for rear wheel drive vehicles often have two separate countershafts, one connected to each input shaft. One example is found in U.S. Pat. No. 5,150,628. These countershafts make the transmission considerably wider than a conventional stepped transmission. That may lead to difficulties in installing the transmission into the vehicle. However, in some dual clutch transmission designs there is only one countershaft, e.g., as in DE923402 and DE3131156A1. On this countershaft there are loose gearwheels arranged that can be rotationally connected to each other and to the countershaft by means of mechanical tooth clutches. In a way, this could be regarded as if the second countershaft is arranged coaxial to the first one. The result will be a power-shiftable dual clutch transmission that is not wider than a corresponding conventional stepped transmission. The number of gears and the corresponding speed reduction ratios are insufficient for heavy severe-duty vehicles, though.
Some dual-clutch transmission concepts in a so-called winding structure have been presented, for instance U.S. Pat. No. 5,347,879, U.S. Pat. No. 5,592,854, DE10325647A1 and DE10339758 A1. In these, the power is led via four gear meshes in at least one gear, and several gearwheels are used for more than one gear. That will give further reduction of speed. However, this corresponds to just one or two additional gears. These concepts are hence less suited to heavy vehicles.
For rear wheel drive, DE102005030987A1, DE102005033027A1, DE 102006015661 A1 and EP1624232A1 show dual clutch transmissions that are connected in series with a range section. That makes them suited for heavy vehicle use. Unfortunately, there are shifts between consecutive gears where the power transfer will be interrupted in these designs.
Two further examples of a dual clutch main transmission being combined with a range section are shown in U.S. Pat. No. 4,966,048 and U.S. Pat. No. 7,204,166B2. These designs comprise two countershafts and will, due to a rear-mounted combined splitter- and range section, have several power interruptions at shifts between consecutive gears. Therefore, they are not favourable from an installation point of view, and not with respect to vehicle performance.
Two similar designs of dual clutch transmission in combination with a range section are shown in DE102005050067A1 and WO2007/039021A1. Therein, the input and output of the transmission can be connected by a friction clutch. This friction clutch can transfer power when a gear shift takes place. Thereby, power interruption can be avoided at all gear shifts. However, for reasonable sizes of this friction clutch, the power transferred to the driven wheels is very small at shifts between low gears. At the same time, the power dissipated in this friction clutch is large at these gear shifts. Thus, these types of dual clutch transmission would have a limited practical use, especially for on- and off-road vehicles.
The dual-clutch transmissions in DE923402 and DE3131156A1 could be combined with a range section. That would give a compact transmission with several gears and high reduction ratios. Gear-shifts between consecutive gears would be without power interruption except when the range section is shifted. This would probably be acceptable on most heavy on-road vehicles, but not for, e.g., trucks in hilly applications or articulated haulers.
U.S. Pat. No. 7,353,724B2 shows in FIGS. 1 and 3 dual-clutch transmissions where a direct connection between one of the input shafts and the output shaft can transmit power when changing between low, underdrive, gears and high, overdrive, gears. This is not a true range section, though. The number of gears is doubled, but in the low, underdrive, gears the power is transmitted via two gear meshes, only, as in FIG. 2 in U.S. Pat. No. 6,958,028B2. That limits the practically possible speed reduction. Thereby, these transmissions are not well suited to heavy vehicles.
A somewhat similar principle is disclosed in U.S. Pat. No. 4,777,837. There, separate gearwheel pairs are provided for intermediate gears between the low and high range gears. This will give a large number of gears and no power interruptions at gear-shifts between consecutive gears. In low range gears, the power is transmitted via three gear meshes, which will enable large reduction ratios. However, the transmission is bulky due to two parallel countershafts. Moreover, the output shaft is not coaxial with the input shaft. That makes the transmission incompatible with most heavy truck designs. The number of components is large, adding costs.
Further on, U.S. Pat. No. 7,070,534B2 presents a dual clutch transmission 10 with a planetary range section 56 and coaxial input 86 and output 68. A dual clutch unit 20, 22 selectively transfers power to input shafts 90 and 92. To each of these input shafts 90 and 92 a countershaft, 74 and 76, respectively, is arranged. From each of these countershafts 74, 76 the power can be selectively directed with tooth clutches 80 and 84 to the output 68 in either of two ways. Firstly, the power can be led to the sun gear 58 of the planetary range section 56 via gearwheels 44, 46 and 54, 46, respectively. That will give a speed reduction in the planetary range section 56, thus corresponding to low range gears. Secondly, the power can be led more directly to the output 68 via gearwheels 40, 42 and 50, 42, respectively. The planetary range section will then be idling, and high range gears are established. Shifts without power interruption can be carried out between low and high range gears. Unfortunately, the number of rotating components, e.g., gearwheels and tooth clutches, is relatively large in comparison with the number of gears. The large number of gearwheels makes the transmission long, and the two parallel countershafts make it wide and difficult to fit in the vehicle. Furthermore, the idling planetary range section will imply unnecessarily large power losses in high range gears.
U.S. Pat. No. 6,958,028B2, FIG. 5, shows a dual clutch transmission with a planetary range section. This transmission is similar to the one in U.S. Pat. No. 7,070,534B2. The main difference is that both input shafts, 30 and 40, use the same countershaft 50, tooth clutch 130, and gearwheels 122, 132 and 132, 128 between this countershaft and the planetary range section. Power interruption between low and high range gears is eliminated by a bridge torque path via a separate countershaft 152. That makes the transmission wide, and it shares the rest of the disadvantages of the one in U.S. Pat. No. 7,070,534B2; many components, long, and high power losses for high range gears.
FIG. 1 in US2008/0188342A1 shows a dual clutch transmission with one countershaft and a planetary range section. A bridge torque path is formed by a tooth clutch 84 that rotationally locks a loose gearwheel 64 on main shaft 28 to a planet carrier 68 rotationally fixed to output shaft 70. When power is led in this path, the gearwheels in the planetary range section are idling, and the range section can be shifted between high and low positions. This gives a narrow transmission with high reduction ratios where power interruptions can be avoided at every shift between consecutive gears. Unfortunately, this bridge path embodiment has many drawbacks. Firstly, the tooth clutch 84 is of complex design, making it costly and long. Secondly, the bearing 32 that carries main shaft 28 must be located in front of loose gearwheel 64. This puts a large part of main shaft 28 behind bearing 32, which, in turn, increases the misalignments in the range section and tooth clutch 84. Moreover, the assembly of the transmission is not facilitated by a main shaft having loose gearwheels and tooth clutches on both sides of the housing wall that carries bearing 32. Thirdly, the addition of parts for the tooth clutch 84 will make the already complex shape of planet carrier 68 even more complex and difficult to produce. DE102007047671A1 shows a similar design that has similar disadvantages.
It is desirable to present an alternative powershift transmission arrangement where the drawbacks of above mentioned prior art are eliminated.
It is desirable to present an improved powershift transmission.
The device according to an aspect of the invention is a powershift transmission in a motor vehicle, where said powershift transmission is arranged between a prime mover and driven wheels of said motor vehicle for transmission of propulsive power and selection of different gear speed ratios, where said powershift transmission comprising at least two frictional clutches for alternatively engaging at least two input shafts, a main transmission, a range section and an output shaft, said main transmission comprising said input shafts, a main shaft and a countershaft, that is parallel to said main shaft, and where said countershaft carries a number of gearwheels that are in mesh with gearwheels that are carried by said main shaft or said at least two input shafts, and where said main shaft is arranged as an output member of said main transmission and integral with or rotationally fixed to an input member of said range section,
and where an output member arranged in said range section is integral with or rotationally fixed to said output shaft, and where said range section is arranged with at least two alternating torque paths with different range speed ratios that can be established by selective engagement and disengagement of at least one range clutch, characterized in that said countershaft is rotationally connectable to said output shaft in order to establish a by-pass torque path, passing by said range section without putting any load on any of the parts in said range section, and in which propulsive power can be transferred when said range section is gear shifted between said different range speed ratios.
The benefit with the device is that for heavy, on- and off-road vehicles the invention provides a transmission that i) enables high power transfer to the driven wheels during all shifts between consecutive gears, ii) can provide high reduction ratios, iii) is cost-effective and simple to produce, iv) has low power losses and v) can be easily installed in a vehicle as an alternative to conventional stepped transmissions.
In another embodiment according to the invention in said by-passing torque path the propulsive power is led between a by-passing gearwheel, which is coaxial with and rotationally connectable to said countershaft, and a by-passing output gearwheel, which is coaxial with and rotationally connectable to said output shaft.
In a further embodiment according to the invention said by-passing gearwheel and said by-passing output gearwheel are in mesh.
In another embodiment according to the invention said by-passing gearwheel is arranged on a by-passing shaft that is integral with or rotationally fixed to said countershaft.
In another embodiment according to the invention said by-passing gearwheel is a loose gearwheel that can be selectively rotationally locked to said by-passing shaft by a by-passing clutch.
In a further embodiment according to the invention said by-passing output gearwheel is a loose gearwheel that can be selectively rotationally locked to said output shaft by a by-passing clutch.
In another embodiment according to the invention said by-passing shaft is arranged to drive a power take-off unit.
In a further embodiment according to the invention said by-passing output gearwheel is located at the same axial position behind said range section as an optional retarder drive gearwheel that can drive an optional retarder auxiliary brake.
In another embodiment according to the invention said by-passing output gearwheel is arranged to drive a retarder.
In a further embodiment according to the invention said by-passing clutch in engaged state rotationally locks said countershaft with said by-passing gearwheel or a by-passing shaft that is integral with or rotationally fixed to said by-passing gearwheel.
In another embodiment according to the invention said range section is of planetary gear design.
In a further embodiment according to the invention said planetary range section comprising a planet carrier that is integral with or rotationally fixed to said output shaft.